US4722186A - Dual pressure displacement control system - Google Patents

Dual pressure displacement control system Download PDF

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Publication number
US4722186A
US4722186A US06/821,947 US82194786A US4722186A US 4722186 A US4722186 A US 4722186A US 82194786 A US82194786 A US 82194786A US 4722186 A US4722186 A US 4722186A
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United States
Prior art keywords
pressure
control
valve
main loop
servo
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Expired - Lifetime
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US06/821,947
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English (en)
Inventor
Joseph E. Louis
Craig C. Klocke
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Danfoss Power Solutions Inc
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Sundstrand Corp
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Priority to US06/821,947 priority Critical patent/US4722186A/en
Assigned to SUNDSTRAND COMPANY, A CORP. OF DE. reassignment SUNDSTRAND COMPANY, A CORP. OF DE. ASSIGNMENT OF ASSIGNORS INTEREST. Assignors: KLOCKE, CRAIG C., LOUIS, JOSEPH E.
Priority to SE8700105A priority patent/SE463828B/sv
Priority to CA000527522A priority patent/CA1273268A/en
Priority to JP62012602A priority patent/JPS62171502A/ja
Priority to DE19873701940 priority patent/DE3701940A1/de
Publication of US4722186A publication Critical patent/US4722186A/en
Application granted granted Critical
Assigned to SUNDSTRAND-SAUER COMPANY, A GENERAL PARTNERSHIP OF DE reassignment SUNDSTRAND-SAUER COMPANY, A GENERAL PARTNERSHIP OF DE ASSIGNMENT OF ASSIGNORS INTEREST. Assignors: SUNDSTRAND CORPORATION, A DE CORP.
Assigned to SAUER INC., reassignment SAUER INC., ASSIGNMENT OF ASSIGNORS INTEREST. Assignors: SUNDSTRAND-SAUER COMPANY, A DE GENERAL PARTNERSHIP
Assigned to SAUER-DANFOSS INC. reassignment SAUER-DANFOSS INC. CHANGE OF NAME (SEE DOCUMENT FOR DETAILS). Assignors: SAUER INC.
Anticipated expiration legal-status Critical
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/002—Hydraulic systems to change the pump delivery
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/42—Control of exclusively fluid gearing hydrostatic involving adjustment of a pump or motor with adjustable output or capacity
    • F16H61/423—Motor capacity control by fluid pressure control means
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/44—Control of exclusively fluid gearing hydrostatic with more than one pump or motor in operation
    • F16H61/456—Control of the balance of torque or speed between pumps or motors

Definitions

  • This invention is directed to a dual source of pressure used in a hydraulic servo control system to modify the displacement of a variable displacement hydraulic unit wherein the pressure for the control system comes from a low pressure source providing normal control of the unit displacement, and an alternate source, preferably one of fluid lines leading from the hydraulic unit, which provides a high pressure source of control fluid.
  • the dual pressure source control system is particularly advantageous for use with a variable displacement bent axis motor encountering stall conditions.
  • variable displacement hydraulic units It is common in variable displacement hydraulic units to have a hydraulic servo control mechanism for varying unit displacement.
  • a low pressure source of control fluid such as a charge pump in a transmission, provides control fluid to the servo mechanism through a displacement control valve.
  • this low pressure source generally in the neighborhood of 200 to 300 psi, provides sufficient fluid pressure to operate the hydraulic servos with proper response speeds under normal operating conditions.
  • variable displacement hydraulic units such as modern design bent axis variable displacement motors which change displacement by sliding a valve plate segment in the endcap, have high internal friction especially in a stall condition when there is no shaft rotation.
  • Such units when stalled require relatively high servo forces to move the valve plate segment. This necessitates either large servo pistons, which are more expensive and increase unit volume, or a source of high pressure. Therefore, some hydraulic units are provided with relatively small servos which are designed to utilize unit main loop working pressure as the control fluid source.
  • high pressure sources are utilized to provide control fluid, unacceptable response speed changes occur relative to variable system working pressures. This is especially true in the case of dual path systems where both transmission motors must react together in a coordinated, controlled manner.
  • a high pressure source for control fluid may require additional or modified pumps and will increase power loss during normal operation.
  • Motor pressure compensating systems have been used wherein main loop working pressures are applied to the motor servos only upon an overpressure condition. These systems do not have any low pressure control fluid source and the displacement is not varied other than in overpressure conditions.
  • Pump pressure limiter systems generally have a displacement control utilizing a low pressure source, such as charge pressure, and sense high working pressure to provide an additional control signal upon an overpressure condition, but such additional control signal is limited by valves to also be of a low pressure nature.
  • Units such as taught in U.S. Pat. No. 4,480,963, Ring issued Nov. 6, 1984, use the highest of main loop pressures and charge pressure. Therefore, the pump servos are always subjected to the highest pressure available and thus cannot have a response rate substantially independent of working pressure. The latter is also true of units such as taught in U.S. Pat. No. 4,194,364, Pahl et al. issued Mar. 25, 1980.
  • the primary feature of the present invention is to provide a dual control pressure source, or a selectively alternative high pressure source of control fluid, for use in a hydraulic unit displacement control servo mechanism wherein the low pressure source is utilized under normal control conditions and the high pressure source is utilized only momentarily under certain temporary conditions.
  • This permits the use of a relatively small volume servo mechanism, reducing overall unit size, with a low pressure source of control fluid, and which is momentarily subjected to a high pressure source when a particular control function is desired.
  • Another object of the present invention is to provide a dual source, or a selectively alternative high pressure source of control fluid as described above, wherein the high pressure source is the working pressure of the hydraulic unit which is connected to the control system by means of a flow restriction orifice so that the control servos are only subjected to the unit high working pressures to initiate servo response while permitting the servo system to have a response rate which is substantially independent of working pressures.
  • the hydraulic unit is provided with main loop lines, at least one of which is capable of being subjected to high loop pressure during operation of the hydraulic unit.
  • a control line including a displacement control valve provides a controlled flow of fluid under pressure to the servo mechanism.
  • One source of fluid under pressure for the control line comprises a low pressure source connected to the control line through a check valve and the high pressure source comprises a high pressure control line connected to the control line downstream of the check valve.
  • the high pressure line includes a flow restriction which limits the flow to the control line and generates a significant pressure drop in the high pressure control line once movement is initiated in the servo mechanism.
  • Still another object of the present invention is to utilize a selectively alternative high pressure control source for a variable displacement hydraulic unit servo mechanism positioning displacement setting means for the unit which is positioned within a hydraulic main loop having at least one main loop line, which during operation of the hydraulic unit is subjected to hydraulic pressure higher than said low pressure source, and is related to hydraulic load.
  • a low pressure source of control fluid is connected to the servo mechanism by a control line including a displacement control valve.
  • the improvement comprises a check valve in the control line between said low pressure source and said displacement control valve and a high pressure source of control fluid consisting of a fluid line connecting the high pressure main loop line to the control line downstream of the check valve and upstream of said displacement control valve.
  • the fluid line includes a normally closed valve which is responsive to main loop pressure which opens when the pressure in the main loop rises above a predetermined pressure.
  • the main loop line provides high pressure to initiate movement of the said servo mechanism when the normally closed valve opens in response to a rise in main loop pressure.
  • a flow restriction can limit the flow from the main loop to the servo mechanism when the normally closed valve is open. The flow restriction provides a significant drop in pressure in the fluid line once the servo mechanism movement is initiated.
  • FIG. 1 is a fragmentary view of a portion of a bent axis hydraulic motor with a displacement control valve and a relatively small volume servo mechanism.
  • FIG. 2 is a schematic view of the hydraulic circuit providing the dual pressure control fluid sources of the present invention.
  • FIG. 2A is a schematic view of an alternative valve which may be used in the control circuit of FIG. 2.
  • FIG. 1 teaches a servo control mechanism for one form of a variable displacement hydraulic unit, and more particularly a bent axis hydraulic motor having a relatively small volume servo mechanism.
  • a motor housing endcap 10 is shown having a curved surface upon which slides a valve plate segment 14 which changes the angle between the hydraulic unit rotating block 16 and an output shaft (not shown). Since the valve plate segment 14 is subjected to the hydraulic unit working pressures, the sliding friction between the segment 14 and the curved surface 12 increase dramatically as hydraulic unit working pressures increase.
  • valve plate segment 14 has a projection 18 which engages a notch 20 of a displacement setting servo piston 22 sliding in bore 24.
  • the projection 18 is representative of many means interconnecting any servo mechanism and a hydraulic unit displacement setting means.
  • the servo piston 22 is subjected to fluid pressures in chambers 26 and 28 which are representative of servo cylinders.
  • a displacement control valve 30 preferably of the proportional type is connected to the two chambers by fluid lines 32 and 34 and is further connected to a control fluid line 36 and to case or drain by drain line 38.
  • the control fluid line 36 is connected to sources of control fluid which are the object of the present invention and are described in detail below.
  • the displacement control valve 30 is positioned by displacement position input means 40, which in the present instance is shown as a spring centered piston arrangement subjected to an input signal ⁇ P.
  • the proportional displacement servo control in the present instance, further includes a feedback spring 42 interposed between the displacement control valve 30 and a wing or projection 44 extending from the servo piston 22.
  • the servo mechanism and displacement control valve means described above are representative of one known form of hydraulic unit variable displacement servo control. While the input shown is a pressure differential input signal, the input signal to the displacement control valve 30 may also be manual, mechanical, or electrical. The above description is merely to disclose one type of hydraulic unit which can utilize the dual pressure control system of the present invention.
  • FIG. 2 shows a pair of variable displacement hydraulic units M R and M L each having the servo mechanism and displacement control valve such as just described.
  • the hydraulic units represent a pair of variable displacement motors utilized in a dual path transmission system.
  • the displacement control servo mechanisms are connected by the control fluid line 36 to a relatively low pressure source 46 of control fluid such as a charge pump.
  • control fluid such as a charge pump.
  • Such servo control is well known whether utilized with a single hydraulic pump or motor, or a pair thereof in dual path systems, with the charge pump providing control pressures in the range of 200-300 psi.
  • the above described servo control provides uniform relatively predictable servo mechanism response rates since the low pressure source 46 is a constant pressure.
  • this control system does not always provide adequate control forces. This is especially true with a bent axis motor unit as shown which has very high internal friction when the motor is stalled. This generates high starting friction, sometimes referred to as stiction in the displacement control and, which must be overcome by the servo mechanism to initiate movement of the valve plate segment 14.
  • the low pressure source 46 is quite often insufficient to provide such servo forces to overcome the stiction, particularly when the aforesaid smaller volume servo mechanism is used.
  • a high pressure fluid control line 50 is connected with control line 36, or lines 36 in a dual path system, at a junction 52 downstream of low pressure source 46 and upstream of the displacement control valve 30.
  • a check valve 54 is interposed between junction 52 and the low pressure source 46 to prevent source 46 from being subjected to high pressures.
  • the high pressure control line 50 is connected to the working pressure of the transmissions as would occur in main loop lines 56 or 58 of the right transmission including the motor M R , and loop lines 56' or 58' connected to the left transmission motor M L .
  • the main loop lines may be connected in open circuit or closed circuit to a transmission pump (not shown).
  • the higher pressure side of the main loop alternates between lines 56 and 58, or lines 56' and 58', dependent upon direction of operation of the transmission.
  • the higher pressure of each transmission main loop is selected by shuttle valve 60 or 60' and the higher of the two transmission pressures is selected by main shuttle 62 which is in turn connected to high pressure control line 50.
  • shuttle valve 60 or 60' is selected by main shuttle 62 which is in turn connected to high pressure control line 50.
  • main shuttle 62 which is in turn connected to high pressure control line 50.
  • shuttles 60' and 62 would not be utilized with the high pressure control line 50 connected directly to shuttle 60.
  • the orifice 64 prevents any substantial bleeding of fluid from the main loop lines, and more importantly generates a pressure drop between line 50 and line 36 once the flow between the high pressure control line 50 and the servo mechanisms has been initiated.
  • the displacement control valves 30 are in their neutral position, there is no flow to the servo mechanisms from either the high pressure line 50 or the low pressure force 46. Movement of the displacement control valves 30 to either their forward or reverse positions permits flow to the servo cylinders.
  • check valve 54 Since line 50 is normally at higher pressure than the low pressure source 46, check valve 54 is closed, and there is high pressure flow from line 50 to the servo cylinders to initiate movement of the servo pistons 22. This initial surge of high pressure fluid overcomes any stiction that might occur in the hydraulic units. Movement of the servo pistons 22 requires flow which generates a pressure drop at orifice 64, which in turn substantially reduces the line 50 pressure at junction 52 until it drops below the pressure of source 46. At this time, check valve 54 opens and further control movement is generated by the regulated 300 psi low pressure source 46.
  • control fluid which could also be a high pressure pump or other source different than the working pressure of the hydraulic units, is only used to initiate servo piston movement, normal control with the predictable response rates is provided from the low pressure source. Even when the high pressure source is the working fluid of the hydraulic unit, the flow from line 50 occurs only momentarily and thus the control response rates are substantially independent of the variable working fluid pressures. It is thus seen that the high pressure control fluid is only provided to the servo mechanism to initiate movement to overcome stiction or when the servo pistons move at extremely low speeds. This is because of the extreme flow restriction generated by orifice 64.
  • the high pressure control line 50 may also be provided with an inline relief valve 66 having upstream and downstream pilots 68 and 70 and set to open at a predetermined pressure by adjustable spring 72.
  • inline relief valve 66 has an inherent characteristic of providing a pressure drop across tne valve. Normally, the maximum pressure of the transmission main loops are determined oy main loop relief valves normally set at an extremely high pressure such as 6000 psi.
  • adjustable spring 72 so as to open the valve 66 at 4500 psi, junction 52 only sees the high pressure when one of the transmission main loops is subjected to pressures above 4500 psi and the valve generates an equal pressure drop.
  • control flow occurs in line 50 only when the inline relief valve 66 senses pressures above 4800 psi which indicate a stall or near-stall condition of one of the motors, M R or M L
  • the hydraulic servo mechanisms are not subjected to any high pressure control source, even to initiate movement of the servo pistons 22, until relatively high working pressures are encountered by the hydraulic units, which is when stiction is highest.
  • High pressure flow to the servo mechanisms is limited to only when the hydraulic unit working pressures are high, thus further making servo response speeds independent of working pressure.
  • the servo mechanisms can be momentarily subjected to as much as 1500 psi control pressures, until there is a flow induced pressure drop at orifice 64, to overcome the stiction in the servo mechanism of the hydraulic units.
  • the flow restriction 64 can be built into, or made an integral part of, the relief valve 66.
  • a pressure responsive sequence valve 66' as shown in FIG. 2A, can be substituted for the inline relief valve 66 in high pressure control line 50.
  • the sequence valve 66' has a high pressure pilot 68' and an adjustable spring 72' similar to the inline pressure relief valve 66 of FIG. 2.
  • the downstream pilot 74 communicates with drain rather than the downstream portion of line 50.
  • the sequence valve 66' does not have the inherent pressure reducing function that occurs with the inline relief valve 66. Therefore, a second flow restricting orifice 76 may also be positioned in the high pressure control line 50 upstream of the sequence valve 66'.
  • the high pressure control fluid flow at junction 52 may be somewhat higher or lower for the embodiment shown in FIG. 2A when compared with the embodiment shown in FIG. 2.
  • the two embodiments work substantially the same with tne flow induced pressure drop in high pressure control line 50 being generated by either one orifice 64 or by a pair of orifices 64 and 76.
  • the flow restrictions 64 and 76 can be built into, or made an integral part of, the sequence valve 66'.
  • the servo mechanisms only see high pressure fluid once the transmission main loops are subjected to a predetermined pressure, and in all cases only see the high pressure fluid momentarily due to the pressure drop at the flow restricting orifice 64, even if there is no valve 66 or 66'. Therefore, response rates substantially independent to working fluid pressure are obtained.
  • the pressure in line 50 available to overcome stiction that is before the flow induced pressure drop, is at least twice the pressure of the low pressure source, at least at maximum pressure or stall conditions.
  • inline valve 66 is used, this is controlled by adjusting spring 72 which adjusts the inherent pressure drop of the valve.
  • the ratio may be as high as 20:1, tnat is 6000 psi (maximum main loop pressure) divided by 300 psi (charge pump pressure). The same is true when a sequence valve is used.
  • both servo mechanisms When the present invention is utilized with a pair of displacement control servo mechanisms, such as in the dual path system shown in FIG. 2, the high pressure flow to both servo mechanisms is only momentary resulting in that both servo mechanisms operate substantially independent of working pressures. Therefore, both servo mechanisms, for all normal control, are subjected to the same low pressure source 46 to operate in a predictable response curve.
  • the highest working pressure available is utilized to overcome servo stiction when necessary in either transmission with both servos being equally subjected to the high pressure to operate in a coordinated manner.
  • Previous dual path systems having low volume servo means had each motor servo means subjected to the higher working pressure in the main loop of each motor resulting in two different response rates, which were further dependent upon working pressures, resulting in inconsistent speed changes for the transmissions.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Control Of Fluid Gearings (AREA)
  • Fluid-Pressure Circuits (AREA)
  • Control Of Transmission Device (AREA)
US06/821,947 1986-01-24 1986-01-24 Dual pressure displacement control system Expired - Lifetime US4722186A (en)

Priority Applications (5)

Application Number Priority Date Filing Date Title
US06/821,947 US4722186A (en) 1986-01-24 1986-01-24 Dual pressure displacement control system
SE8700105A SE463828B (sv) 1986-01-24 1987-01-14 Deplacementreglersystem med tvaa tryck
CA000527522A CA1273268A (en) 1986-01-24 1987-01-16 Dual pressure displacement control system
DE19873701940 DE3701940A1 (de) 1986-01-24 1987-01-23 Wahlweise alternative steuerdruckversorgung fuer eine hydraulische verstelleinheit
JP62012602A JPS62171502A (ja) 1986-01-24 1987-01-23 二重圧力置換制御システム

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
US06/821,947 US4722186A (en) 1986-01-24 1986-01-24 Dual pressure displacement control system

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US4722186A true US4722186A (en) 1988-02-02

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US06/821,947 Expired - Lifetime US4722186A (en) 1986-01-24 1986-01-24 Dual pressure displacement control system

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US (1) US4722186A (enrdf_load_stackoverflow)
JP (1) JPS62171502A (enrdf_load_stackoverflow)
CA (1) CA1273268A (enrdf_load_stackoverflow)
DE (1) DE3701940A1 (enrdf_load_stackoverflow)
SE (1) SE463828B (enrdf_load_stackoverflow)

Cited By (11)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US5191950A (en) * 1990-05-22 1993-03-09 Linde Aktiengesellschaft Hydrostatic travelling mechanism for track-laying vehicles
US5245828A (en) * 1989-08-21 1993-09-21 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for civil engineering and construction machine
WO2001027472A1 (de) * 1999-10-12 2001-04-19 Brueninghaus Hydromatik Gmbh Verstellvorrichtung einer schrägscheibenkolbenmaschine
WO2003095831A1 (de) * 2002-05-10 2003-11-20 Brueninghaus Hydromatik Gmbh Verstellvorrichtung für hydrostatische kolbenmaschinen
US20040083728A1 (en) * 2001-06-13 2004-05-06 Juan Moya Hydrostatic drive
US20090301076A1 (en) * 2008-06-06 2009-12-10 Toshifumi Yasuda Vehicle Transaxle System
WO2012171206A1 (zh) * 2011-06-16 2012-12-20 长沙中联重工科技发展股份有限公司 变量泵的恒功率控制装置、方法以及混凝土泵送装置
WO2014026788A1 (de) * 2012-08-17 2014-02-20 Robert Bosch Gmbh Aktoreinrichtung und axialkolbenmaschine
WO2015066182A1 (en) * 2013-10-29 2015-05-07 Raven Industries, Inc. Hydraulic displacement control system
US9549504B2 (en) 2008-06-06 2017-01-24 Kanzaki Kokyukoki Mfg. Co., Ltd. Transaxle system for vehicle
US20230105578A1 (en) * 2021-10-04 2023-04-06 Hamilton Sundstrand Corporation Variable positive displacement pump actuator systems

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102007044451A1 (de) * 2007-09-18 2009-03-19 Robert Bosch Gmbh Anschlussplatte für eine hydrostatische Kolbenmaschine
DE102017213458A1 (de) * 2017-08-03 2019-02-07 Robert Bosch Gmbh Hydrostatische Axialkolbenmaschine mit Leistungsbegrenzung

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US3499285A (en) * 1967-07-21 1970-03-10 Dowty Hydraulic Unit Ltd Hydraulic apparatus
US4376371A (en) * 1979-10-09 1983-03-15 Kabushiki Kaisha Komatsu Seisakusho Hydraulic circuit for a hydraulically driven vehicle
US4480963A (en) * 1982-11-22 1984-11-06 Deere & Company Pump swashplate control assist
US4627239A (en) * 1980-06-06 1986-12-09 Kawasaki Jukogyo Kabushiki Kaisha Hydraulic circuit arrangement

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US3908519A (en) * 1974-10-16 1975-09-30 Abex Corp Control systems for a variable displacement pump
DE2723074A1 (de) * 1977-05-21 1978-11-30 Bosch Gmbh Robert Verdraengermaschine mit einem druckregler
US4476680A (en) * 1979-08-14 1984-10-16 Sundstrand Corporation Pressure override control
US4512723A (en) * 1983-10-17 1985-04-23 Sundstrand Corporation Pressure limiter
DE3340332C2 (de) * 1983-11-08 1988-11-10 Hydromatik GmbH, 7915 Elchingen Leistungs-Regelvorrichtung für einen hydrostatischen Antrieb mit Fördermengeneinstellung

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US3499285A (en) * 1967-07-21 1970-03-10 Dowty Hydraulic Unit Ltd Hydraulic apparatus
US4376371A (en) * 1979-10-09 1983-03-15 Kabushiki Kaisha Komatsu Seisakusho Hydraulic circuit for a hydraulically driven vehicle
US4627239A (en) * 1980-06-06 1986-12-09 Kawasaki Jukogyo Kabushiki Kaisha Hydraulic circuit arrangement
US4480963A (en) * 1982-11-22 1984-11-06 Deere & Company Pump swashplate control assist

Cited By (19)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US5245828A (en) * 1989-08-21 1993-09-21 Hitachi Construction Machinery Co., Ltd. Hydraulic drive system for civil engineering and construction machine
US5191950A (en) * 1990-05-22 1993-03-09 Linde Aktiengesellschaft Hydrostatic travelling mechanism for track-laying vehicles
WO2001027472A1 (de) * 1999-10-12 2001-04-19 Brueninghaus Hydromatik Gmbh Verstellvorrichtung einer schrägscheibenkolbenmaschine
US6725658B1 (en) 1999-10-12 2004-04-27 Brueninghaus Hydromatik Gmbh Adjusting device of a swashplate piston engine
US7163078B2 (en) * 2001-06-13 2007-01-16 Brueninghaus Hydromatik Gmbh Hydrostatic drive
US20040083728A1 (en) * 2001-06-13 2004-05-06 Juan Moya Hydrostatic drive
US7171887B2 (en) 2002-05-10 2007-02-06 Brueninghaus Hydromatik Gmbh Regulating device for hydrostatic piston machines
US20050252369A1 (en) * 2002-05-10 2005-11-17 Winfried Lilla Regulating device for hydrostatic piston machines
WO2003095831A1 (de) * 2002-05-10 2003-11-20 Brueninghaus Hydromatik Gmbh Verstellvorrichtung für hydrostatische kolbenmaschinen
DE10220889C5 (de) * 2002-05-10 2009-05-28 Brueninghaus Hydromatik Gmbh Verstellvorrichtung für hydrostatische Kolbenmaschinen
US20090301076A1 (en) * 2008-06-06 2009-12-10 Toshifumi Yasuda Vehicle Transaxle System
US9211793B2 (en) * 2008-06-06 2015-12-15 Kanzaki Kokyukoki Mfg. Co., Ltd. Vehicle transaxle system
US9549504B2 (en) 2008-06-06 2017-01-24 Kanzaki Kokyukoki Mfg. Co., Ltd. Transaxle system for vehicle
WO2012171206A1 (zh) * 2011-06-16 2012-12-20 长沙中联重工科技发展股份有限公司 变量泵的恒功率控制装置、方法以及混凝土泵送装置
WO2014026788A1 (de) * 2012-08-17 2014-02-20 Robert Bosch Gmbh Aktoreinrichtung und axialkolbenmaschine
WO2015066182A1 (en) * 2013-10-29 2015-05-07 Raven Industries, Inc. Hydraulic displacement control system
US10087960B2 (en) 2013-10-29 2018-10-02 Raven Industries, Inc. Hydraulic displacement control system
US20230105578A1 (en) * 2021-10-04 2023-04-06 Hamilton Sundstrand Corporation Variable positive displacement pump actuator systems
US11994117B2 (en) * 2021-10-04 2024-05-28 Hamilton Sundstrand Corporation Variable positive displacement pump actuator systems

Also Published As

Publication number Publication date
DE3701940A1 (de) 1987-07-30
DE3701940C2 (enrdf_load_stackoverflow) 1992-10-08
SE463828B (sv) 1991-01-28
JPS62171502A (ja) 1987-07-28
SE8700105D0 (sv) 1987-01-14
SE8700105L (sv) 1987-07-25
CA1273268A (en) 1990-08-28

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